1. Field of the Invention
The present invention relates generally to a fuel control system for a cylinder injection type internal combustion engine (also known as the direct fuel injection type engine) for a motor vehicle in which fuel is injected directly into engine cylinders to control engine output torque. More particularly, the present invention is concerned with a fuel control system for the cylinder injection type internal combustion engine which system is so designed as to ensure brake operating pressure while protecting positively combustion performance or combustibility of the engine against degradation.
2. Description of Related Art
In general, in the internal combustion engine employed as the automobile engine or the like, an injector for fuel injection is installed within an intake manifold of an intake pipe of the engine so that the fuel as injected can be charged into engine cylinders together with the intake air.
For having better understanding of the principle underlying the invention, technical background thereof will be described in some detail. FIG. 10 is a schematic diagram showing a conventional fuel control system for an internal combustion engine in which a fuel injector is disposed within an intake pipe.
Referring to FIG. 10, an engine 1 constituting a main body of the internal combustion engine system includes a plurality of cylinders. However, for simplification of the illustration, only one of the cylinders is representatively shown in FIG. 10.
An intake pipe 1a is communicated to an exhaust pipe 1b through combustion chambers of the engine 1 to which a crank shaft 1c is coupled at one end thereof.
The intake pipe or manifold 1a serves for charging a mixture of intake air and fuel (hereinafter also referred to as the air-fuel mixture) into the engine 1, while the exhaust pipe 1b is used for discharging the exhaust gas resulting from the combustion of the air-fuel mixture within the engine 1. The crank shaft 1c is driven rotationally by the engine 1. Cooling water 1d forced to flow around the engine 1 serves for cooling the engine 1.
An air flow sensor 2 installed at an inlet port of the intake pipe 1a measures intake air quantity Qa as the information concerning the air flow rate or amount of air supplied to the engine 1. Furthermore, mounted within the intake pipe 1a is a throttle valve 3 which is operatively coupled to an accelerator pedal (not shown) operated by a driver of the motor vehicle for regulating the intake air quantity Qa in dependence on the magnitude of depression stroke of the accelerator pedal.
For the purpose of detecting angular position of the throttle valve 3, i.e., throttle opening degree .theta. of the throttle valve 3, a throttle position sensor 4 is provided in association with the throttle valve 3.
Further provided in association with the crank shaft 1c is a crank angle sensor 5 which is designed for detecting rotation speed (rpm) of the crank shaft 1c to thereby output a pulse signal in synchronism with the rotation of the crank shaft 1c. This signal will be referred to as the crank angle signal SGT. Thus, the crank angle signal SGT carries the information concerning the rotation speed (rpm) of the engine 1 as well as the information concerning the angular position of the crank shaft 1c (i.e., crank angle).
Temperature Tw of the cooling water 1d is detected by a water temperature sensor 6 which thus can serve as a means for detecting a warmed-up state of the engine 1.
An O.sub.2 -sensor 7 provided in association with the exhaust pipe 1b is designed to detect an oxygen concentration Do of the exhaust gas discharged from the engine 1 to the exhaust pipe 1b.
For the purpose of controlling operations of the internal combustion engine system described above, a control unit 8 is provided, which may be implemented in the form of a microprocessor or microcomputer. The detection signals Qa, .theta., SGT, Tw and Do outputted from the various sensors 2, 4, 5, 6 and 7 installed at the peripheral portions of the engine 1, as mentioned above, are supplied as input information signals to the control unit 8 which in turn outputs driving control signals for various devices and actuators such as spark plugs and fuel injectors (described hereinafter) in dependence on the operation states of the engine to thereby perform various sequential controls inclusive of the ignition timing control and fuel injection control for each of the cylinders of the engine 1. Owing to such arrangement as mentioned above, the engine 1 can be driven through combustion of the air-fuel mixture at the desired ignition timing with the desired air-fuel ratio.
A spark plug 9 is mounted within each of the cylinders of the engine 1, being exposed to the combustion chamber defined within the cylinder, wherein the firing of the spark plug 9 is controlled by an ignition timing control signal P outputted from the control unit 8.
As can be seen in FIG. 10, a bypass passage BP is provided in parallel to the intake pipe 1a so that the intake air can controllably bypass the throttle valve 3.
More specifically, operation of the air bypass valve 10 provided at the bypass passage BP is controlled by a bypass control signal B outputted from the control unit 8, whereby the rate of bypass air flow (i.e., bypassed intake air quantity) Qb which bypasses the throttle valve 3 can be regulated by selectively opening or closing the bypass passage BP. In this manner, there can be realized not only the engine torque control in the running state of the motor vehicle but also the engine rotation speed control in the idling operation state of the engine (the engine operation state in which the throttle valve 3 is fully closed).
Referring continuously to FIG. 10, a fuel injector 11 is mounted within the intake manifold located downstream of the intake pipe 1a. Operation of the fuel injector 11 is controlled by the fuel injection control signal J outputted from the control unit 8, whereby controlled quantity of the fuel is supplied to the engine cylinders.
An exhaust gas recirculation pipe (hereinafter also referred to as the EGR pipe) EP through which the exhaust pipe 1b is communicated with the intake pipe 1a serves for recirculating the exhaust gas discharged from the engine 1 into the combustion chamber thereof with a view to reducing harmful components of the exhaust gas such as nitrogen oxides or NO.sub.x by burning again the exhaust gas.
An exhaust gas recirculation valve 12 (hereinafter also referred to as the EGR valve) mounted on the EGR pipe EP is driven by an EGR control signal E issued from the control unit 8 to thereby control the amount or quantity of the exhaust gas (referred to as the EGR quantity in short) recirculated from the exhaust pipe 1b to the intake pipe 1a.
A cylinder identifying sensor 13 mounted on the cam shaft of the engine 1 outputs to the control unit 8 a cylinder identifying signal SGC for identifying the combustion cylinder in synchronism with the operation of the intake valve of the engine 1.
A check valve 15 communicated to the intake manifold of the intake pipe 1a serves to hold a lower limit value of an intake or manifold pressure (negative pressure or vacuum) Pi within the intake pipe 1a as a brake operating pressure PB.
Communicated to the check valve 15 is a master bag 16 which serves for storing the brake operating pressure PB in a predetermined negative pressure state or at a predetermined vacuum level)
The check valve 15 and the master bag 16 cooperate to constitute a brake operating pressure generating means which may also be referred to as the brake pressure multiplication mechanism for generating a negative brake operating pressure PB on the basis of the intake pressure (intake-manifold pressure) Pi of the engine 1. By using the brake operating pressure PB as a driving energy source, a braking mechanism (not shown) is actuated for assisting the operator or driver of the motor vehicle in his or her brake manipulation.
More specifically, when the operator or driver detaches his or her foot from the accelerator pedal, the throttle valve 3 is closed, as a result of which the intake air quantity fed to the cylinders of the engine 1 decreases, whereby the intake pressure Pi is stored as a negative-pressure source for the brake force to be used for actuating the braking mechanism (hereinafter referred to simply as the brake).
As is apparent from the above, the intake pressure Pi of the engine 1 is used as the brake operating pressure PB for actuating the brake, wherein the intake pressure Pi is stored in the master bag 16 by way of the check valve 15 in the engine operation state such as the idling operation, brake applying operation or the like in which the intake pressure Pi becomes negative (e.g. the intake pressure Pi assumes a minimum value PO, as will be described hereinafter).
On the other hand, in the engine operation state such as the ordinary running state in which the intake pressure Pi is high, leakage of the negative pressure from the master bag 16 to the intake pipe is prevented owing to the action of the interposed check valve 15. In this manner, the brake operating pressure PB (negative pressure or vacuum) is prevented from consumption so long as the brake is not applied. Thus, the brake operating pressure PB is ordinarily held at a pressure level or vacuum level which does not exceed the intake pressure Pi.
The detection signals Qa, .theta., SGT, Tw, Do and SGC derived from the outputs of the air flow sensor 2, the throttle position sensor 4, the crank angle sensor 5, the water temperature sensor 6, the O.sub.2 -sensor 7 and the cylinder identifying sensor 13, respectively, are inputted to the control unit 8. On the other hand, the various components or devices such as the spark plug 9, the air bypass valve 10, the fuel injector 11 and the EGR valve 12 are driven in response to control signals P, B, J and E, respectively, which are outputted from the control unit 8.
In the conventional fuel control system for the indirect injection type engine implemented in such arrangement as shown in FIG. 10, when the fuel injection control signal J is outputted from the control unit 8, the fuel injector 11 is driven in response thereto during a period corresponding to the pulse width of the fuel injection control signal J, whereby a quantity of fuel indicated by the fuel injection control signal J is injected into the intake pipe 1a.
However, when the fuel is injected by the injector mounted externally of the engine cylinder, as shown in FIG. 10, a part of the fuel will adhere to inner walls of the intake pipe 1a and surfaces of the intake valves of the engine 1 before the fuel is actually charged into the engine cylinder. In this conjunction, it is noted that such fuel deposition is likely to occur particularly when the engine is operating at a low temperature (such as engine starting operation in which the fuel is relatively difficult to vaporize) or when the engine is in a transient operation state where the amount of fuel to be supplied to the engine has to be changed at a high response speed. Thus, reduction of the contents of harmful gas components carried by the exhaust gas is subjected to adverse influence.
Under the circumstances, there has been proposed a cylinder injection type internal combustion engine system which is designed for injecting the fuel directly into the engine cylinders with a view to solving the problem mentioned above.
Such direct fuel injection engine system (or cylinder injection type engine system) attracts attention as an ideal engine system. When the above-mentioned engine system is adopted in place of the conventional gasoline engines for the motor vehicle, there can be realized very advantageous effects, which will be mentioned below.
(1) Reduction of Harmful Gas Components Contained in the Exhaust Gas
Since the fuel is directly injected into the combustion chamber defined within the engine cylinder in the vicinity of the spark plug 9 (refer to FIG. 10), the air-fuel ratio may be increased so that the air-fuel mixture becomes lean without taking into consideration the delay involved in the transportation of the fuel, whereby contents of harmful HC (hydro carbon) gas and CO (carbon monoxide) gas carried by the exhaust gas can be reduced.
(2) Reduction of Fuel Consumption
Because the fuel is injected immediately before the ignition timing under the control of the ignition timing signal, there is formed a mass of combustible fuel mixture around the spark plug 9 at the time of ignition, rendering nonuniform the distribution of the gas mixture containing the fuel. Thus, the fuel-air mixture undergoes a so-called stratified combustion. By virtue of this feature, the air-fuel ratio in appearance between the amount of air and that of the fuel charged into the engine cylinder can be significantly increased with the air-fuel mixture being made lean correspondingly.
Besides, owing to realization of the stratified combustion mentioned above, combustion of the air-fuel mixture is less affected adversely even when the exhaust gas is recirculated with an increased ratio (i.e., notwithstanding of increased exhaust gas recirculation or EGR in abbreviation). Besides, the intake air quantity Qa can be increased, which in turn means that so-called pumping loss can be reduced. For these reasons, the fuel-cost performance of the engine can be enhanced significantly.
(3) Increased Output Power of the Engine
Since the air-fuel mixture tends to be concentrated around the spark plug 9, the amount of end gas (air-fuel mixture gas in the regions located remotely from the spark plug 9) decreases favorably owing to the effects of the stratified combustion mentioned previously, whereby the anti-knocking performance of the engine can be enhanced with the compression ratio in the engine 1 being significantly increased.
Furthermore, because the fuel is converted into gas or gasified within the cylinder, the intake air is deprived of heat as vaporization heat. Consequently, the density of the intake air can be increased, which is effective for enhancing the volumetric efficiency.
(4) Enhancement of Drivability
By virtue of the system of injecting directly the fuel into the engine cylinder, the time taken for generating the output torque of the engine 1 through the fuel combustion, starting from the fuel injection, can be shortened when compared with the indirect injection type engine system shown in FIG. 10. Thus, the internal combustion engine system of the cylinder injection type can respond to the demand of the driver with high speed.
Parenthetically, in the fuel control system for the cylinder injection type internal combustion engine, there exists two combustion modes, i.e., a so-called lean mode (or lean operation mode) in which a little amount of fuel is supplied during the compression stroke for thereby enhancing the emission performance as well as the fuel-consumption performance of the engine owing to excessively lean or over-lean stratified combustion and a so-called stoichiometric mode (rich operation mode) in which a predetermined or stoichiometric quantity of fuel is supplied to the engine during the suction stroke with a view to increasing the engine output torque by realizing ordinary uniform mixture combustion.
In the compression-stroke fuel injection mode (lean operation mode), the engine is operated with a lean air-fuel mixture when compared with that in the suction-stroke fuel injection mode (stoichiometric or rich mode), it is required to charge into the engine 1 an increased amount of intake air Qa for a given throttle opening degree .theta. (a given depression stroke of the accelerator pedal). Consequently, the intake air quantity Qa controlled ordinarily only with the accelerator pedal by the driver has to be controlled by resorting to another control means in order to increase the intake air quantity Qa.
Now, description will be made of a conventional cylinder injection type internal combustion engine system. FIG. 11 is a schematic diagram showing generally a structure of a conventional fuel control system for a cylinder injection type internal combustion engine such as disclosed in Japanese Unexamined Patent Application Publication No. 186034/1987 (JP-A-62-186034). In the figure, components like as or equivalent to those described hereinbefore by reference to FIG. 10 are designated by like reference characters and repeated description in detail of these components is omitted.
In the cylinder injection type internal combustion engine now of concern, the fuel control system is so designed as to adjust or regulate supplementarily the fuel injection quantity in response to change or variation of the fuel pressure in order to suppress the change in the engine output torque.
Referring to FIG. 11, the control unit 8A is so designed as to determine arithmetically the fuel supply quantity and the fuel injection timing for outputting the fuel injection control signal J in accordance with the result of the arithmetic operation, to thereby allow a fuel injector 11A to be driven during at least one of the suction stroke and the compression stroke for injecting the fuel. In that case, by identifying the control-subjected cylinder on the basis of the cylinder identifying signal SGC, the fuel injector 11A can be controlled on a cylinder-by-cylinder basis.
The fuel injector 11A is not mounted within the intake pipe 1a but installed within the cylinder so as to be directly exposed to the combustion chamber of the engine 1 and is implemented with high-speed/high-pressure specifications so that the fuel of high pressure can be injected into the cylinder within a short time period during the suction or compression stroke.
A fuel injector driver 14 interposed between the control unit 8A and the fuel injector 11A to drive the fuel injector 11 serves to convert a fuel injection control signal J issued by the control unit 8A to a fuel injection control signal K for the high-speed/high-pressure fuel injection to thereby drive the fuel injector 11A.
Since the injector driver 14 is designed so as to output a fuel injection control signal K of a large electric power by amplifying the power of the fuel injection control signal J issued by the control unit 8A, as mentioned above, the fuel can be injected at a pressure sufficiently high for overcoming the pressure prevailing within the cylinder.
Further, the air bypass valve 10A of the engine system now under consideration is so arranged as to control the bypass intake air quantity Qb over a wide range when compared with the air bypass valve 10 of the engine system described hereinbefore by reference to FIG. 10 in order to control the engine output torque during engine operation in the lean mode (i.e., mode in which the engine operates with combustion of lean air-fuel mixture) inclusive of cruising mode in addition to the control of the engine rotation number in the idling operation mode in which the throttle valve 3 is fully closed.
At this juncture, comparison of the cylinder injection type internal combustion engine system shown in FIG. 11 with the system described hereinbefore by reference to FIG. 10 shows that the former differs structurally from the latter in that the fuel injector 11A for supplying the fuel is not mounted within the intake pipe 1a but installed directly within the cylinder of the engine 1 and is implemented with high-speed/high-pressure specifications so that the fuel can be injected into the cylinder at a high pressure within a short time period during the suction or the compression stroke, and that the injector driver 14 is additionally provided to drive the fuel injector 11A of the high-speed/high-pressure specifications.
In the following, operation of the conventional fuel control system for the internal combustion engine of the cylinder injection type (or direct fuel injection type) shown in FIG. 11 will be elucidated.
It is noted that in the fuel control system for the cylinder injection type engine, the fuel control is performed in a stratified combustion mode (i.e., combustion mode in which excessively lean or over-lean air fuel mixture is burnt) by supplying the fuel to the engine cylinder immediately before the ignition timing (the compression-stroke injection). Accordingly, the air-fuel ratio A/F is so controlled as to be more than 30 inclusive, falling within the over-lean range. It should however be noted that the air-fuel ratio of the mixture actually undergoing combustion is close to the stoichiometric air-fuel ratio A/F of 14.7.
In the case of the conventional indirect injection type internal combustion engine shown in FIG. 10, combustion takes place at the air-fuel ratio A/F of ca. 20 (lean burn) in the suction-stroke injection mode after the intake air and the fuel have been mixed uniformly. Differing from the engine system shown in FIG. 10, in the cylinder injection type internal combustion engine shown in FIG. 11, combustion takes place at the air-fuel ratio A/F of ca. 16 (at which nitrogen oxides (NO.sub.x) are produced at high rate). Such being the circumstances, in the cylinder injection type internal combustion engine, a large amount or quantity of exhaust gas is recirculated to the engine with a view to realizing reduction of nitrogen oxides (NO.sub.x) contained in the exhaust gas discharged from the engine.
As will now be understood from the foregoing, in the cylinder injection type internal combustion engine shown in FIG. 11, lean combustion in the compression-stroke injection is carried out through combination of the stratified combustion which can be realized by subtle control of the fuel injection timing and the ignition timing and the recirculation of a large amount or quantity of exhaust gas which may lead to degradation of the combustion performance of the engine 1.
On the other hand, in the control mode in which high output torque of the engine is demanded as in the case of acceleration of the motor vehicle, the stoichiometric combustion (i.e., combustion of rich air-fuel mixture) is performed in the suction-stroke injection mode, similarly to the indirect injection type engine shown in FIG. 10, for thereby realizing combustion of uniform mixture.
In general, when the engine operation state is changed over from the over-lean combustion state in the compression-stroke injection mode to the rich combustion state in the suction-stroke injection mode, not only the air-fuel ratio and the EGR quantity but also the fuel injection timing and the ignition timing are supplementarily adjusted or regulated. In that case, the intake air quantity Qa is decreased by controlling correspondingly the air bypass valve 10A in order to prevent the output torque of the engine from fluctuation which may otherwise be brought about upon changeover of the fuel injection mode from the compression-stroke injection mode (lean combustion mode) to the suction-stroke injection mode (rich combustion mode).
For changing over the fuel injection mode from the compression-stroke injection mode (lean mode) to the suction-stroke injection mode (rich mode), a large number of control parameters have to be changed substantially concurrently in order to afford change of the combustion state.
In conjunction with the concurrent changeover of many control parameters, it is however noted that because of nonuniformity in the performance among the components subjected to the control, age changing thereof, variation in the environmental conditions during the running of the motor vehicle and/or difference of the combustion states, there may arise such situation in which the combustion state can not make transition smoothly to the combustion of uniform mixture from the stratified combustion, incurring possibly unstable combustion and hence fluctuation of the rotation speed and eventually vibration of the engine 1 upon changeover of the fuel injection control mode.
On the other hand, when the engine rotation number (rpm) becomes lowered due to external load applied to the engine 1 in the lean mode where the fuel supplied in the compression stroke undergoes over-lean stratified combustion, fluctuation of rotation can be suppressed by decreasing the air-fuel ratio A/F (i.e., enriching the air-fuel mixture), because the output power of the engine 1 bears correlation to the fuel supply quantity.
In the field of the conventional internal combustion engines, it has been proposed to suppress the fluctuation of rotation by adjusting supplementarily the ignition timing. However, in the compression-stroke injection mode, the combustion state will change because the phase relation between the ignition timing and the fuel injection timing changes when the ignition timing is altered. Thus, because of difficulty encountered in the control of the ignition timing, as mentioned above, suppression of the fluctuation of the rotation is attempted by adjusting or controlling the fuel supply.
FIG. 12 is a flow chart for illustrating control mode determining operation of the conventional fuel control system for the cylinder injection type internal combustion engine shown in FIG. 11. Further, FIGS. 13 and 14 are views for illustrating arithmetic operations for determining control parameters, i.e., the air-fuel ratio and the ignition timing by mapping in the control mode, i.e., the lean mode and the rich mode, respectively. Incidentally, in FIGS. 13 and 14, engine rotation number Ne (rpm) is taken along the abscissa with engine output torque Te being taken along the ordinate.
More specifically, FIG. 13 is a view for illustrating arithmetic determination of the parameters in the lean mode, i.e., air-fuel ratio A/F1 and ignition timing Pt1 by mapping for an engine rotation number Ne1 (rpm) and an engine torque Te1 in the lean mode. By contrast, FIG. 14 is a view for illustrating arithmetic determination of the parameters in the stoichiometric mode, i.e., air-fuel ratio A/F2 and ignition timing Pt2 by mapping for an engine rotation number Ne2 (rpm) and an engine torque Te2 in the stoichiometric mode (rich mode).
The processing procedure illustrated in FIG. 12 is executed in synchronism with the crank angle signal SGT (indicating a predetermined angular position of the crank shaft or CA in short), as is obvious to those skilled in the art.
Ordinarily, the control unit 8A fetches as the engine operating state information the various sensor signals such as the throttle opening degree .theta., crank angle signal SGT, intake air quantity Qa and the water temperature Tw, to thereby set the control mode (i.e., rich mode or stoichiometric mode) on the basis of the engine operating state and determine the parameters (i.e., air-fuel ratio and ignition timing) for the control mode as set.
Referring to FIG. 12 showing the processing procedure during the ordinary operation of the engine, the control unit 8A first executes the control mode decision processing for deciding whether or not the accelerator pedal is in the released state (also referred to as the off-state or simply as off) in which the accelerator pedal is not depressed (step S1).
When it is decided in the step S1 that the accelerator pedal is in the off-state (i.e., when the decision step S1 results in affirmation or "YES"), it is then decided whether or not the current control mode is a stoichiometric mode set forcibly (step S2). In case the decision step S2 results in negation or "NO", then a decision step S3 is executed to decide whether or not the brake is applied (i.e., the brake pedal is depressed or on).
When it is decided in the step S3 that the brake is applied (i.e., when the decision step S3 results in "YES"), then the stoichiometric mode is activated with a forced stoichiometric mode flag being set (step S4), whereupon the processing proceeds to a control quantity calculation processing step S7.
On the other hand, when it is decided in the step S1 that the accelerator pedal is applied (on-state), i.e., when the decision step S1 results in "NO", the forced stoichiometric mode flag is cleared (step S5), whereon the control mode decision processing (step S6) is executed and then the processing proceeds to the control quantity calculation processing (step S7).
By contrast, when it is decided in the step S2 that the current mode is the stoichiometric mode set forcibly (i.e., when the decision step S2 results in "YES"), then the stoichiometric mode activating processing and the forced stoichiometric mode flag setting processing are executed (step S4) without executing the brake application (on) decision step S3.
Furthermore, when it is decided in the step S3 that the brake is off (i.e., when the decision step S3 results in "NO"), then the control mode decision processing is executed in the step S6.
In the control quantity calculation processing (step S7), decision is first made as to whether or not the current control mode is stoichiometric mode (step S8).
Unless the current mode is decided as the stoichiometric mode (i.e., when the decision step S8 results in "NO"), then the current mode is regarded as the lean mode. Accordingly, the air-fuel ratio A/F1 and the ignition timing Pt1 for the lean mode (compression-stroke injection mode) are arithmetically determined in a step S9, as illustrated in FIG. 13, whereupon the processing exits the procedure shown in FIG. 12.
At this juncture, it should be added that in the step S9, the ignition timing Pt1 and the fuel injection timing for the lean mode are arithmetically determined and additionally the EGR valve 12 installed in the EGR pipe EP is opened to allow an EGR quantity EGR1 to be introduced into the intake pipe 1a from the exhaust pipe 1b while the air bypass valve 10A installed in the bypass passage BP is opened to allow a bypass intake air quantity Qb1 to increase considerably. This is for the purpose of increasing the air-fuel ratio A/F to thereby make lean the air-fuel mixture.
On the other hand, when it is decided in the decision step S8 that the current mode is the stoichiometric mode (i.e., when the decision step S8 results in "YES"), the air-fuel ratio A/F2 and the ignition timing Pt2 for the stoichiometric mode (suction-stroke injection mode) are arithmetically determined in a step S10, as illustrated in FIG. 14, whereupon the processing exits the routine shown in FIG. 12.
Further, it should be added that in the step S10, the ignition timing Pt2 and the fuel injection timing for the stoichiometric mode are arithmetically determined and at the same time an EGR valve 12 installed in the EGR pipe EP is closed to allow the exhaust gas recirculation (EGR) from the exhaust pipe 1b to the intake pipe 1a to be interrupted or cut while the air bypass valve 10A is closed to allow the bypass intake air flow to be interrupted or cut. This is for the purpose of decreasing the air-fuel ratio A/F to thereby make rich the air-fuel mixture.
FIG. 15 is a timing chart for illustrating concretely typical operation of the conventional fuel control system for the cylinder injection type internal combustion engine shown in FIG. 11. In FIG. 15, time t is taken along the abscissa. As can be seen in the figure, it is presumed that the accelerator pedal is actuated (ON) at a time point t1, starting from the initial idling state, and the brake application is repetitively activated (ON) at time points t2, . . . , t5, respectively.
Referring to FIG. 15, the brake operating pressure PB is held at a minimum value PO in the idling state where the throttle valve 3 (refer to FIG. 11) is fully closed. In this state, the intake air quantity supplied to the engine 1 is controlled through the bypass passage.
Immediately after the time point t1 when the accelerator pedal is actuated or depressed, starting from the idling state, the stoichiometric control mode is selected in order to ensure a sufficiently large output torque of the engine 1. In that case, the intake pressure Pi (indicated by a single-dot curve) increases steeply. Nevertheless, the brake operating pressure PB represented by a solid-line curve is maintained at the minimum value PO under the action of the check valve 15.
Further, when the intake pressure Pi increases to a level Pi1 (pressure level in the stoichiometric mode in the brake-released state) to assume an equilibrium state, after the accelerator pedal has been depressed. Then, the control mode is changed over to the lean mode from the stoichiometric mode. In that case, the intake pressure Pi increases further to a level Pi2 (the lean-mode pressure in the brake-released or -off state) to assume the equilibrium state due to increase of the intake air quantity in the lean mode.
Subsequently, when the accelerator pedal is released with the brake being applied at the time point t2, the intake pressure Pi decreases to the level Pi3 (lean-mode pressure in the brake-on state), whereas the brake operating pressure PB increases (i.e., vacuum level of the brake operating pressure decreases, to say in another way) due to consumption of the brake operating pressure PB.
More specifically, every time the brake is applied at the time points t2, t3 and t4, respectively, the vacuum level of the brake operating pressure PB rises from the minimum value PO to the values PB3, PB4 and then to the value PB5 progressively in this order. At a time point t5 immediately after the time point t4 at which the brake operating pressure PB exceeds the upper limit value (threshold value) PTH which delimits the range enabling the brake operating pressure PB to be active, the brake operating pressure PB becomes equal to the intake pressure Pi3, rendering unavailable the assist efforts for the brake application.
By the way, in the compression-stroke injection mode, the engine operates with lean air-fuel mixture when compared with the engine operation in the suction-stroke injection mode. Accordingly, in the compression-stroke injection mode, it is required to increase the intake air quantity fed to the engine 1 even when the operator actuates the throttle valve to a same opening degree as in the suction-stroke injection mode by actuating the accelerator pedal, as a result of which the intake pressure (intake manifold pressure) Pi becomes high in the compression-stroke injection mode.
In the state where the intake pressure is high, as mentioned above, intake air loss (pumping loss) of the engine 1 is low with the load torque imposed on the engine 1 being decreased, which is favorable from the standpoint of the fuel-cost performance of the engine.
In this conjunction, it is noted that in the cylinder injection type engine, the engine can operate in the over-lean operation state, as mentioned previously. Accordingly, even when the operator detaches his or her foot from the accelerator pedal for the brake operation, engine operation in the lean operation state can be continued without reducing the intake air quantity, as a result of which the fuel-consumption performance of the engine can be improved.
It is however noted in conjunction with the engine operation in the lean mode that such situation is likely to occur in which sufficiently high brake operating pressure PB (brake power source of negative pressure or vacuum) is not available. In particular, in case the brake operating pressure PB stored in the master bag 16 has been consumed due to the so-called pumping brake operation, there may be incurred degradation in the brake performance.
The problem mentioned above may naturally take place even in a Diesel engine which includes no throttle valve and in which the intake pressure of negative level can not exist inherently. However, in the case of the Diesel engine, the above problem is solved by employing a vacuum pump.
By contrast, in the case of the cylinder injection type engine, the intake air quantity is controlled for the over-lean stratified combustion by supplying the fuel in the compression stroke (lean operation mode), as described hereinbefore. Consequently, difficulty is encountered in ensuring the intake pressure Pi of a vacuum level during the engine operation in the lean mode.
Certainly, it is conceivable that when the intake pressure of a vacuum level is required for the brake control, the intake air quantity is decreased by supplying the fuel during the suction stroke to thereby allow the ordinary uniform mixture combustion (stoichiometric combustion) to take place with a view to ensuring the intake pressure of a vacuum level.
FIG. 16 is a timing chart for illustrating operations involved in changing over the control mode from the lean mode to the stoichiometric mode every time the brake is applied at the time points t2, t3, t4 and t5, respectively.
As can be seen in FIG. 16, the brake operating pressure PB converges to the pressure PS of a vacuum level in the stoichiometric mode which is slightly higher than the minimum level PO in the idling mode. Further, the level of the pressure PS in the stoichiometric mode becomes lower as the rotation speed (rpm) of the engine 1 increases.
Referring to FIG. 16, the control mode of the engine 1 is changed over to the stoichiometric mode from the lean mode at the time points t2, t3, t4 and t5, respectively.
In that case, the bypass intake air flow Qb is cut in the stoichiometric mode (see step S10 shown in FIG. 12), which results in that the intake air quantity Qa fed to the engine 1 decreases. Thus, the negative pressure or vacuum for the brake operating pressure PB can be ensured.
Consequently, the intake pressure Pi is held at a level not higher than the intake pressure Pi2, as indicated by a single-dot broken curve shown in FIG. 16, while the brake operating pressure PB (indicated by a solid line curve) is held at a level not higher than the upper limit value PB11. Thus, the brake effort assisting capability can be protected against degradation.
However, assurance of the negative pressure or vacuum by changing over the operation mode from the compression-stroke injection (lean operation state) to the suction-stroke injection (stoichiometric operation state) and repetition of the compression-stroke injection (lean operation) and the suction-stroke injection (stoichiometric operation) are unfavorable from the standpoint of the combustion performance, involving ultimately deterioration of the drivability of the motor vehicle.
As will now be understood from the foregoing description, the conventional fuel control system for the cylinder injection type internal combustion engine suffers a problem that the brake operating pressure PB increases in the lean mode, incurring degradation in the brake performance because no consideration is paid to the assurance of the brake operating pressure PB (negative pressure or vacuum).
Furthermore, when the engine operation mode is changed over to the ordinary stoichiometric mode (operation with rich air-fuel mixture) in the course of the lean operation (i.e., operation with lean air-fuel mixture) with the aim of ensuring the brake operating pressure PB, then the compression-stroke injection and the suction-stroke injection are alternately repeated at a high frequency, degrading the combustibility or combustion performance, which in turn leads to degradation of the drivability, giving rise to another problem.